High speed rotor assembly

ABSTRACT

A ring-shaped magnet (19) is bonded on or secured to the outer peripheral side of a rotor (21) which is borne by a rotary shaft (13) which constitutes a radial pneumatic dynamic pressure bearing. A ring-shaped magnet (23) is disposed so that it is spaced at a given distance from the ring-shaped magnet (19). The rotor (21) is restricted from moving in the axial direction and is rotatably supported in the radial direction by the magnetic attracting forces of two ring-shaped magnets (19, 23). The rotor (21) is eccentric by 0.5 relative to the axis of rotation by the difference in magnetic balance between the ring-shaped magnets (19, 23). Accordingly, the radial pneumatic dynamic pressure bearing generates a high dynamic pressure. The herringbone pneumatic dynamic pressure generating grooves (15a, 15b) generate a high dynamic pressure. They provide a very high radial load bearing ability to enhance the accuracy of rotation.

This application is the national phase of international applicationPCT/JP95/01078, filed Jun. 2, 1995 which designated the U.S.

This application is the national phase of international applicationPCT/JP95/01078, filed Jun. 2, 1995 which designated the U.S.

BACKGROUND OF THE INVENTION

1. FIELD OF INVENTION

The present invention relates to a high speed rotor assembly that isstrongly required to stably rotate at a high accuracy, such as rotarybearing for polygon mirror.

2. RELATED ART

Polygon mirrors are used in, for example, laser printers, bar codereaders, facsimiles and laser copy machines. Since the efficiency of thepolygon mirror is enhanced as the number of rotations increases, thenumber of rotations of 20000 r.p.m. or more has been preferred . As thebearing for such a high speed rotor assembly, for example, precise ballbearings have heretofore been used. However, it is difficult to keep astable accuracy of rotation for an extended period of time due to suchproblems as wearing and that the limit of the rotational speed is about16000 r.p.m.

Accordingly, grooved pneumatic dynamic pressure bearings for bearing arotor rotating at a high speed have been used. The grooved pneumaticdynamic pressure bearings have features of low noise and low frictionloss since they bear a rotor via a gas film in a non-contacting manner.However, they are weak for external shocks and vibrations of the rotorare may occur since the load bearing ability of the bearing is low. Thusit is difficult to obtain a stable accuracy of rotation.

In order to overcome the above-mentioned problems, there has beenproposed, inter alia, a high speed rotor assembly in which the bearingrigidity is increased to suppress the vibration of the rotor 52 byenlarging the bearing 51 as disclosed in Japanese Unexamined PatentPublication No. Tokkai Sho 60-98213 and shown in the sectional view ofFIG. 8. Another proposed solution has been a high speed rotor assembly,as disclosed in Japanese Unexamined Patent Publication No. Tokkai Sho63-266420 and as shown in the section view of FIG. 9, in which: thrustdirection pneumatic dynamic pressure bearings 62a and 62b are providedat opposite ends of the radial pneumatic dynamic pressure bearing 61 tosuppress the vibrations of the rotor 63 by the pneumatic dynamicpressure generated in a thrust direction; and the size of a rotary shaft61 is reduced.

In these high speed rotor assemblies which use axially long pneumaticdynamic pressure bearing having an increased radial rigidity for bearingthe rotor , the rotor is larger than the rotor in the high speed rotorassembly using conventional ball bearings, thereby resulting in anincrease in the size of the assemblies. One the other hand, thepneumatic dynamic pressure bearings generally require precise machiningin which the accuracy of shape such as degree of cylindricality,roundness and surface roughness should be in the order of 1 μm toprovide excellent accuracy of rotation. Therefore, the fact that therotary shaft is long in an axial direction makes it difficult to performmachining. Although it is possible to perform machining, themanufacturing cost is so high that it is impractical.

Since the area of contact between the rotor and the thrust bearingsurface is large in the rotor assembly in which the radial pneumaticdynamic pressure bearing is provided at both ends thereof with thethrust dynamic pressure bearings, a high starting torque is required forstarting the rotor. As a result ringing effect (adsorption effect on thesmooth face) mayoccur on the thrust bearing face which may preventactivation of the rotor assembly.

When the inner peripheral side of a cylindrical work piece is machined,the inner diameter at opposite ends tends be less or higher than that atthe midposition thereof. In other words, the inner peripheral side ofthe work piece may be machined into arcuate shape as shown in FIGS. 2Aand 2B. Therefore, unstable dynamic pressure distribution is liable tooccur at the both ends of the radial bearing due to the changes in thegap of the bearing and the resulting changes in the number of rotations.Vibrations in a conical mode may occur due to occurrence of unforeseenvibrations, amplification of the vibrations, of the rotor at a specificvibration frequency of synchronized vibration and resonation of thevibration with the magnetic force balance in a thrust direction. As aresult, the bearing of the rotor may contact with the stationary shaftand the vibrations due to shock may not be dumped.

Therefore, the present invention was made in order to overcome theabove-mentioned problems. It is an object of the present invention toprovide a high speed rotor assembly which is compact in size; requires aless starting torque; and provides a superior stable accuracy ofrotation in a wide range from the beginning of rotation to a high speedrotation to a high shock environment.

SUMMARY OF THE INVENTION Means for Accomplishing the Above MentionedObject

According to an aspect of the present invention, there is provided ahigh speed rotor assembly including:

a radial direction pneumatic dynamic pressure bearing having a housing,a stationary shaft which erects on and within said housing and a rotaryshaft surrounding said stationary shaft, which is rotatably borne; and

a magnetic bearing having a rotor which is secured to said rotary shaftand which is provided along its outer peripheral side with a magnetizedportion, a ring-shaped magnet secured to said housing, which ispositioned in-the same plane as said magnetized portion for bearing saidrotor in a radial and thrust directions so that a gap is formed betweensaid rotary and stationary shafts in a radial direction,

characterized in that the radial load of said rotor is borne by thecombination of the load bearing ability of said radial pneumatic dynamicpressure bearing and the radial load bearing ability of said magneticbearing and in that said rotor is borne at an eccentricity of 0.3 to0.7. A term "eccentricity" used herein is defined as the offset of thecenter of the rotor from the center of the radial bearing/the gap ofradial bearing.

The load bearing ability of said radial pneumatic dynamic pressurebearing for bearing said rotor may be larger than the radial loadbearing ability of said magnetic bearing in a range of applicablenumbers of rotations of said rotor (3000-40000 r.p.m.).

The bearing surfaces of said stationary and rotary shafts in said radialpneumatic bearing may be formed into smooth surfaces.

Said stationary shaft of said radial pneumatic bearing may be formed onthe bearing surface thereof with herringbone pneumatic dynamic pressuregenerating grooves and annular grooves which is continuous to the bothends of said herringbone pneumatic dynamic pressure generating grooves.

Said annular grooves formed on the bearing surface of said stationaryshaft may have a depth which is larger than that of said herringbonepneumatic dynamic pressure generating grooves.

Said annular grooves may extend beyond the edge of the radial bearingsurface of said rotary shaft by 1/3 to 2/3 of its width outwardly in anaxial direction.

Said stationary and rotary shafts may be made of ceramic material.

According to another aspect of the present invention, there is provideda high speed rotor assembly including:

a radial pneumatic dynamic pressure generating grooves, a stationaryshaft which erects on and within a housing, and a rotary shaftsurrounding said stationary shaft, which is rotatably borne on saidhousing; and

a rotor secured to said rotary shaft,

characterized in that said assembly comprises:

a magnetized portion provided along the outer peripheral side of saidrotor; and

a magnetic bearing having a ring-shaped magnet which is in the sameplane as said magnetized portion for bearing said rotor in a radial andthrust directions so that a gap is formed in a radial direction betweensaid rotor shaft and said stationary shaft and in that the gap betweenthe magnetized portion provided along the outer periphery of said rotoris larger than that between said rotary and stationary shafts.

The gap between said magnetized portion and said ringshaped magnet maybe 0.05 to 5 mm and the gap between said stationary shaft and saidrotary shafts is 1 to 30 μm.

Said magnetized portion may be formed on the outer peripheral side ofsaid rotor and on a plane including the gravity center of said rotor andwhich is normal to the axis of said rotor.

Said stationary shaft of said radial pneumatic bearing may be formed onthe bearing surface thereof with herringbone pneumatic dynamic pressuregenerating grooves.

Said stationary shaft of said radial pneumatic bearing may be formed atthe ends of the bearing surface thereof with annular grooves.

Said bearing surfaces of said stationary and rotary shafts in saidradial pneumatic bearing may be formed as smooth surfaces.

A term "high speed rotor assembly" used herein generally refer to motorshaving a dynamic pressure bearing mechanism.

When the rotor is stopped to be rotated by the magnetic bearing, themovement in an axial direction is restricted by the magnetic force sothat the rotor is levitated in a given position in a thrust direction.As pneumatic dynamic pressure is generated in a gap between thestationary shaft and the rotary shaft by the herringbone grooves, thenumber of rotations of the rotor increases. The rotor will have anincreased bearing rigidity and dumping characteristics in both radialand thrust directions against the external shock over a wide range ofenvironments from an initial rotation to high speed rotation owing tothe combined effect of the pneumatic dynamic pressure with the magneticforce of the ring-shaped magnets. The load bearing ability of thepneumatic dynamic pressure due to the herringbone grooves will becomelarger than that of the magnetic forces from the magnets as the numberof the rotations of the rotor increases. This suppresses the whirlmotion of the rotor (precession of the rotor) to provide a higheraccuracy of rotation at a higher speed.

By forming the bearing surfaces of said stationary and rotary shaftsinto smooth surfaces in the radial pneumatic bearing, the radialpneumatic bearing can be manufactured at a low cost.

By forming the herringbone pneumatic dynamic pressure generating grooveson the bearing surface of the stationary shaft in the radial bearing,the stability can be enhanced.

The annular grooves at the ends of the herringbone grooves formed on thebearing surface of said stationary shaft which is continuous thereto andhas a larger depth prevent the distribution of dynamic pressure whichoccurs at the ends of the rotor.

By making the stationary and rotary shafts of ceramic material, deadlock due to contact between stationary and rotary shafts made of metal,which otherwise occurs, can be prevented.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a sectional view showing a high speed rotor assembly of afirst embodiment of the present invention;

FIG. 2A and 2B is a schematic view for illustrating the shape of arotary shaft of the present invention;

FIG. 3A and 3B is a schematic view for explaining the principle of apneumatic dynamic pressure bearing;

FIG. 4 is schematic view for explaining the whirl movement which occursin the rotor assembly using the pneumatic dynamic pressure bearing;

FIG. 5 is a table showing a result of a simulation in which theperformance of the prior art is compared with a high speed rotorassembly of the present invention;

FIG. 6 is a sectional view showing a high speed rotor assembly of asecond embodiment of the present invention;

FIG. 7 is a table showing a result of simulation of the performance ofthe high speed rotor assembly of the second embodiment;

FIG. 8 is a sectional view of a high speed rotor assembly forillustrating the prior art; and

Fig. 9 is a sectional view of a high speed rotor assembly on which aminiaturized polygon mirror is mounted for illustrating the prior art.

DETAILED DESCRIPTION OF THE INVENTION

Before the description of the presently preferred embodiments of thepresent invention, the principle of a pneumatic dynamic pressure bearingand whirl motion will be described.

Principle of the Penumatic Dynamic Pressure Bearing

The pneumatic dynamic pressure bearing makes use of the characteristicwhich gas possesses as an viscous fluid. As shown in FIG. 3A and 3B,load bearing capability is obtained by two effects. A first effect whichis shown in FIG. 3A is a dynamic pressure effect in which when a gas isentrained into a tapered gap 33 between opposite two surfaces 31 and 23by the sliding movement thereof, a pressure of the gas is built up tobear a depressing force, that is, a load, so that a gas film is formedbetween the sliding surfaces. A second effect which is shown in FIG. 3Bis caused when opposite two surface move toward each other at a givenspeed. At this time, the gas 36 between two surfaces 34 and 35 should beexpelled at a rate which is proportional to the approaching speed of thetwo surfaces 34 and 35. This expelling of the gas causes the pressure ofthe gas to increase against the depressing forces of the two surfaces.This effect is a dynamic pressure effect due to non-steady gas filmformation and is also referred to as squeeze effect.

Principle of Occurrence of Whirl Motion

FIG. 4 is a schematic view explaining the whirl motion of a radialbearing. For ease of explanation, a vertical axis is assumed and noeffect of gravity is considered. If it is assumed that the center of arotary shaft 41 which is rotating at an angular velocity ω is offsetfrom the bearing center 43 by e, a pressure is generated in the gaslubricating film between a fixed bearing 42 and the rotary shaft 41 dueto the above-mentioned dynamic pressure effect. A force Fe which servesto return the rotary shaft 41 to the original bearing center 43 is notonly generated, but also a force F ω which causes precession of therotary shaft 41 around the bearing center 43. When the rotary shaft 41begins precession at an angular speed Ω, a pressure is generated due tothe squeeze effect to cause a force F Ω which serves to suppress theprecession of the shaft 41.

The centrifugal force acting on the rotor 41 is represented by meΩ²where in the mass m of the rotor 41 which is borne by the bearing. Therotor 41 will perform the precession at such an angular velocity Ω thatmeΩ² =Fe. If FΩ=Fω at this time, there is not force which accelerates ordecelerates the precession. Accordingly, the steady precession of therotor with a given radius 44 will continue. If FΩ<Fω, the precession isdecelerated so that the centrifugal force will gradually decrease. IfFΩ>Fω, the precession motion will be accelerated so that the centrifugalforce will surpass the centripetal force and gradually increase.Therefore, in order to suppress the whirl motion from occurring, thecentripetal force component Fe of the reaction of the gas film should beincreased. It is necessary to decrease the influence of the dynamicpressure on the changing component of the reaction of the gas filmwithout decreasing the squeeze effect. In the first embodiment of thepresent invention, the whirl motion is suppressed by increasing thecentripetal component Fe of the reaction of the gas film by making therotary shaft eccentric within a given range with a dynamic pressurecaused by the herringbone shaped grooves and a magnetic force from amagnetic bearing.

Now, the first embodiment of the present invention will be describedwith reference to drawings. FIG. 1 shows the schematic sectionalstructure of a high speed rotor assembly having a polygon mirror of thefirst embodiment. A stationary shaft 12 made of a ceramic material isprovided in the center of a housing 11 made of aluminum so thatstationary shaft is provided on the housing. The ceramic stationaryshaft 12 is formed with dynamic pressure generating grooves, orherringbone dynamic pressure generating grooves 15a, 15b, having a depthof 6 to 8 microns, which are inclined in opposite directions withrespect to the axial direction so that a maximum pressure is generatedin the center 14 of bearing surface of the ceramic rotary shaft 13.Annular grooves 16a and 16b having a depth of 10 microns are formed atthe opposite ends of the herringbone dynamic pressure generating grooves15a, 15b. The annular grooves 16a, 16b have a width of 2.0 mm and extendoutwardly beyond the edge of the bearing surface of the rotary shaft 13by one thirds of their width.

A ring-shaped magnet 23 is bonded on or secured to the inner side of thehousing 11 so that it is coaxial with the ceramic stationary shaft 12. Aflange 17 which is made of aluminum is secured to the outer periphery ofthe ceramic rotary shaft 13 by shrinkage fit. The aluminum flange 17 isformed with recesses on the lower and outer sides thereof. Ring-shapedmagnets 18 and 19 are fitted on the recesses. A polygon mirror 20 issecured to the upper side of the aluminum flange 17 by means of adhesiveor bolts. These components constitute a rotor 21.

The rotor 21 is fitted on the ceramic stationary shaft 12 so that asmall gap or clearance 22 which is in the order of about 5 microns isformed between the inner side of the rotor 21 (rotary shaft 13) and theouter side of the stationary shaft 12 and a gap which is 1.5 mm isformed between the ring-shaped magnet 19 fitted on the outer side of therotor 21 and the inner side of the housing 11. The rotor 21 is preventedfrom moving in an axial direction by the magnetic attracting forces oftwo ring-shaped magnets 19 and 23 and is levitated in a given positionin a thrust direction. Two ring-shaped magnets 19 and 23 are positionedsubstantially coaxially with respect to the center of the bearing whilethe rotor 21 is eccentric relative to the center of the bearing by 2.5μm (eccentricity =0.5) due to the fact that the magnetic balance isslightly different between the magnets 19 and 23. Accordingly, a largedynamic pressure can be generated even at not higher than 10000 r.p.m.and shock resistance can be enhanced.

It is preferable that the gaps 24 and 22 between the magnetized portionalong the outer side of the rotor and the ring-shaped magnet and betweenthe rotor and the ceramic bearing be 0.05 to 0.5 mm and 1 to 30 μm,respectively.

If the gap 24 is not larger than 0.05 mm, the magnetic attracting forceexceeds the radial load bearing capability.

If the gap 24 exceeds 5 mm, the rotor can not be borne in a thrustdirection and the centrifugal force becomes smaller. Accordingly, if thegap is not within this range, it is hard to obtain the eccentricity of0.3 to 0.7 on rotation at a high speed.

If the gap 22 is not larger than 1 μm, the load bearing ability becomestoo high, resulting in unstable rotation.

The friction loss on steady rotation is high and the power consumptionbecomes high. The amount of generated heat becomes high so that variousfunctions including rotation stability and damage due to contact occurby the changes: in dimension of mechanical parts due to thermalexpansion.

If the gap 22 is larger than 30 μm, the load bearing ability is lowered,resulting also in a less rotation stability.

By adjusting the gaps 22 and 24 so that they fall within these ranges,the eccentricity can be stably kept to 0.3 to 0.7 on high speed rotationby the combination of pneumatic bearing and magnetic bearing.

In the present invention, it is preferable that the magnetized portionbe formed on the outer peripheral side of the rotor and on a planeincluding the gravity of the rotor, which is normal to the axis of therotor.

This is the reason why the conical vibrations can be suppressed by theexistence of the magnetized portion on the plane including the gravitycenter.

The rotor 21 is rotated by means of a motor comprising a coil 25 securedto the housing 11 and a magnet 18 secured to the aluminum flange 17 anda rotation control unit (not shown).

The rotor 21 is started to rotate by the motor. As the number ofrotations of the rotor 21 increases, a pressure is generated in thebearing gap 22 by the herringbone pneumatic dynamic pressure generatinggrooves 15a, 15b so that the radial load bearing ability is obtained.The increase in the radial load bearing ability enhances the shockresistance of the bearing and suppresses the whirl motion of the rotor21 so that the radius of the whirl motion becomes small to provideincreased precision of the rotation.

In other words, a pneumatic dynamic pressure is generated in a gapbetween the stationary shaft 12 and the rotary shaft 13 by theherringbone grooves 15a, 15b as the number of rotations of the rotor 21increases. The rotor 21 will have an increased bearing rigidity anddumping characteristics in both radial and thrust directions against theexternal shock over a wide range of environments from an initialrotation to high speed rotation due to the combined effect of thepneumatic dynamic pressure with the magnetic force of the ring-shapedmagnets 19, 23. The load bearing ability of the pneumatic dynamicpressure due to the herringbone grooves 15a and 15b will become largerthan that of the magnetic forces from the magnets 19 and 23 as thenumber of the rotations of the rotor 21 increases. This suppresses thewhirl motion of the rotor to provide a higher accuracy of rotation at ahigher speed.

Unstable dynamic pressure distribution which occurs at the both ends ofthe rotary shaft 13 is suppressed, and contact between both ends of therotary shaft 13 and the stationary shaft 12 is prevented from occurringdue to the fact that the annular grooves 16a, 16b, which are continuouswith the herringbone grooves 15a, 15b, has a larger depth than that ofthe herringbone grooves 15a, 15b and that and extend outwardly beyondthe edge of surface of the radial bearing of the rotary shaft 13 by onethirds of its width. It is considered that the unstable dynamic pressuredistribution can be suppressed by the fact that the annular grooves 16a,16b are offset outwardly by 1/3 to 2/3 of their width relative to theedge of the radial bearing surface of the rotary shaft 13.

By making the stationary and rotary shafts of ceramic material, deadlock due to contact between stationary and rotary shafts made of metalwhich otherwise occurs can be prevented.

In the above-mentioned first embodiment of the high speed rotorassembly, the rotor 21 is eccentric with respect to the center of thebearing at 0.5 by the difference in magnetic balance between thering-shaped magnets 19 and 23 as mentioned above. By making the rotaryshaft 13 eccentric, a higher dynamic pressure is generated to enhancethe accuracy of the rotation. It is found from the result of simulationthat enhancement in the precision of the rotation due to theeccentricity of the rotor 21 provides a remarkable result when theeccentricity is not less than 0.3 on steady rotation. On the other hand,when the eccentricity is 0.7 or higher, the risk of contact between thestationary shaft 12 and the rotary shaft 13 occurs on application of aexternal force such as shock, etc. Therefore, the eccentricity ispreferably 0.3 to 0.7.

The results of comparison between the prior art shown in FIG. 9 and thesimulation of the precision of rotation of the high speed polygon formedas mentioned above is rotated at 30000 r.p.m. is shown in FIG. 5.

In the high speed rotor assembly of the first embodiment, the rotor 21is levitated in an axial direction by the attracting forces of thering-shaped magnets 19 and 23 as mentioned above. Therefore, thefriction loss which occurs before levitation is eliminated so that therise-up time until steady rotation is shorter and less starting torqueis necessary.

Particularly, in the first embodiment, a higher dynamic pressure can begenerated since the rotor 21 is eccentric relative to the center of thebearing by the difference in magnetic balance between the magnets 19 and23. A higher dynamic pressure can also be generated by the herringbonepneumatic dynamic pressure generating grooves 15a, 15b. It is deemedthat these effects combine to generate very high load bearing abilityand to enhance the shock resistance of the bearing. It is furtherconsidered that the whirl motion of the rotor 21 is suppressed so thatthe radius of the whirl motion becomes small to enhance the accuracy ofthe rotation.

It is considered that the mechanism of thrust dynamic pressure bearingis not necessary to reduce the power consumption since the rotor 21 islevitated in an axial direction by the above-mentioned magnets 19 and23. It is understood from FIG. 5 that the high speed rotor assembly ofthe first embodiment has very excellent rotation characteristics.

Now, a high speed rotor assembly of the second embodiment of the presentinvention will be described with reference to FIG. 6. Parts which aresubstantially identical with those of the first embodiment are denotedby like reference numerals and description thereof is omitted.

In the second embodiment, the stationary shaft 112 of ceramic materialis not provided on the surface thereof with herringbone pneumaticdynamic pressure generating grooves 15a, 15b, but its surface issmoothed. Similarly with the first embodiment, the rotor 21 is eccentricrelative to the center of the bearing by 2.5 μm (eccentricity =0.5) bythe difference in magnetic balance between the ring-shaped magnets 19and 23.

In other words, by making the rotary shaft 13 eccentric, a high dynamicpressure is generated to prevent the rotor 21 from performing the whirlmotion to provide a high precision of rotation.

The result of simulation of the performance of the high speed rotorassembly of the second embodiment will be described with reference toFIG. 7. FIG. 7 shows the bearing number, stability characteristics,stability limit weight (kg) for the radial clearances of 0.0020, 0.0025,00030 (mm) when the number of rotations is set to 8000 (r.p.m.) and theeccentricity is 0.1 to 0.8. It is found from the simulation result thata high stability limit weight can be obtained by making the eccentricity0.3 or more. On the other hand, there occurs the risk of contact betweenthe stationary shaft 12 and the rotary shaft 13 due to shocks, etc.,similarly with the above-mentioned first embodiment when theeccentricity is 0.7 or more. Accordingly, it is preferable that theeccentricity be 0.3 to 0.7.

The second embodiment has an advantage in that a high speed rotarypolygon mirror can be provided at a low cost by omitting the herringbonepneumatic dynamic pressure grooves 15a, 15b which are difficult to beprovided on the surface of the stationary shaft 112 of the ceramicmaterial.

As mentioned above, the high speed rotor assembly of the presentinvention can be used for a bearing for a polygon mirror used in thelaser printer, bar code readers, facsimiles and laser copy machines. Therotor is also suitable for recording apparatus, machine tools, measuringinstruments in which a member which is rotating at a high speed is borneby a thrust pneumatic dynamic pressure bearing.

What is claimed is:
 1. A high speed rotor assembly comprising:a radialpneumatic dynamic pressure bearing which includes a housing, astationary shaft provided on and within said housing, and a rotary shaftsurrounding said stationary shaft, said rotary shaft being rotatablyborne; and a magnetic bearing which includes a rotor secured to saidrotary shaft and provided along its outer peripheral side with amagnetized portion, and a ring-shaped magnet secured to said housingpositioned in the same plane as said magnetized portion for bearing saidrotor in radial and thrust directions so that a gap is formed betweensaid rotary and stationary shafts in a radial direction, wherein theradial load of said rotor is borne by the combination of the loadbearing ability of said radial pneumatic dynamic pressure bearing andthe radial load bearing ability of said magnetic bearing, and said rotoris borne at an eccentricity of 0.3 to 0.7.
 2. A high speed rotorassembly as defined in claim 1 in which the load bearing ability of saidradial pneumatic dynamic pressure bearing for bearing said rotor islarger than the radial load bearing ability of said magnetic bearing ina range of applicable numbers of rotations of said rotor.
 3. A highspeed rotor assembly as defined in claim 1 or 2 in which the bearingsurfaces of said stationary and rotary shafts in said radial pneumaticbearing are formed into smooth surfaces.
 4. A high speed rotor assemblyas defined in claim 1 or 2 wherein said stationary shaft of said radialpneumatic bearing is formed on the bearing surface thereof withherringbone pneumatic dynamic pressure generating grooves and annulargrooves, said annular grooves being formed at opposite ends of saidherringbone pneumatic dynamic pressure generating grooves.
 5. A highspeed rotor assembly as defined in claim 4 in which said annular groovesformed on the bearing surface of said stationary shaft have a depthwhich is greater than that of said herringbone pneumatic dynamicpressure generating grooves.
 6. A high speed rotor assembly as claim 5in which said annular grooves extend beyond the edge of the radialbearing surface of said rotary shaft by 1/3 to 2/3 of its widthoutwardly in an axial direction.
 7. A high speed rotor assembly asdefined in claim 4 in which said annular grooves extend beyond the edgeof the radial bearing surface of said rotary shaft by 1/3 to 2/3 of itswidth outwardly in an axial direction.
 8. A high speed rotor assembly asdefined in claim 1 in which said stationary and rotary shafts are madeof ceramic material.
 9. A high speed rotor assembly comprising:a radialpneumatic dynamic pressure bearing which includes a housing,a stationaryshaft provided on and within said housing, and a rotary shaftsurrounding said stationary shaft, said rotary shaft being rotatablyborne; and a magnetic bearing which includes a rotor secured to saidrotary shaft and provided along its outer peripheral side with amagnetized portion, and a ring-shaped magnet secured to said housingpositioned in the same plane as said magnetized portion for bearing saidrotor in radial and thrust directions so that a gap is formed betweensaid rotary and stationary shafts in a radial direction, wherein the gapbetween said magnetized portion and said ring-shaped magnet is 0.05 to 5mm and the gap between said stationary shaft and said rotary shaft is 1to 30 μm.
 10. A high speed rotor assembly as defined in claim 9 whereinsaid magnetized portion is formed on a plane which includes the centerof gravity of said rotor, said plane being normal to the axis of saidrotor.
 11. A high speed rotor assembly as defined in claim 9 in whichsaid stationary shaft of said radial pneumatic bearing is formed on thebearing surface thereof with herringbone pneumatic dynamic pressuregenerating grooves.
 12. A high speed rotor assembly as defined in claim9 in which said stationary shaft of said radial pneumatic bearing isformed at the ends of the bearing surface thereof with annular grooves.13. A high speed rotor assembly as defined in claim 9 in which thebearing surfaces of said stationary and rotary shafts in said radialpneumatic bearing are formed into smooth surfaces.